Thermal refrigeration apparatus



Oct. 29, 1968 c TURNBLADE ET AL 3,407,626

THERMAL REFRIGERATION APPARATUS Filed Feb. 24, 1955 2 Sheets-Sheet 1 5 IyvEA/raeg, fl/mmepif wvezfloe,

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Oct. 29, 1968 Filed Feb. 24, 1965 R. C. TURNBLADE ET AL THERMAL REFRIGERAT ION APPARATUS 2 Sheets-Sheet 2 United States Patent 3,407,626 THERMAL REFRIGERATION APPARATUS Richard C. Turnblade, Northridge, and Oded E. Sturman, Arleta, Califi, assiguors to .Conductron Corporation, Ann Arbor, Mich., a corporation of Delaware Filed Feb. 24, 1965, Ser. No. 434,954 8 Claims. (Cl. 62-498) This invention relates to a refrigeration apparatus and more particularly to an improved refrigeration apparatus which includes a thermally actuated fluid ovi g apparatus.

For many applications the need has arisen for a refrigeration' apparatus which is more compact and with less power requirements than the systems presently known to the art. Additionally, special applications of refrigeration systems to unusual environmental conditions of zerogravity, acceleration, vibration, and orientation such as encountered in aerospace utilization have given rise to refrigeration and air conditioning problems not easily and efficiently solved by systems and apparatus known to the prior art. Under such conditions conventional apparatus, such as typical piston or reciprocating pumps, for moving the fluid through a refrigerating system are not suitable due to the weight, electrical power requirements and number of moving parts which are subject to wear. The necessity for extreme reliability and long life in such uses is apparent. In order to avoid the electrical power requirements attempts to employ pumping and compressing apparatus which utilizes available thermal energy has been attempted. Thermal pumps and compressors known to the prior art have not, in general, been satisfactory due to their complexity and relatively low efficiency.

Accordingly, it is an object of the present invention to provide an improved refrigeration apparatus which can be operated by thermal energy as the prime energy source.

It is another object of the present invention to provide an improved refrigerating apparatus which is extremely reliable and has a long operating life.

It is another object of the present invention to provide a refrigerating apparatus which is of minimal size and weight.

Another object of the present invention is to provide a refrigerating apparatus which can be operated by waste heat which is the product of equipment in connection with which the refrigerating apparatus is used.

It is another object of the present invention to provide a thermally actuated refrigerating apparatus having modest heat power requirements.

A primary object of the present invention is to provide a refrigerating apparatus in which pumping fluid and refrigerating fluid can be intermixed or a single fluid used for both functions.

In general, the present invention comprises a refrigerating apparatus in which the heat extracted from a heat source is used to vaporize a pumping fluid contained within a closed chamber. The heat available is sufficient to boil or vaporize the fluid within the chamber to increase the internal pressure within the chamber. When the internal pressure reaches a predetermined level an outlet valve opens and vaporization of the liquid within the chamber continues under a constant pressure process. The high pressure fluid can be utilized to drive a refrigerant fluid compressor in 'a single fluid embodiment in which compressor the fluid is transposed to the required pressure and temperature to complete the condensation expansion steps in a normal refrigeration cycle. In the two fluid embodiment compatible pumping fluid and refrigerant fluids are intermixed such that the admixed fluids provide the compression-condensation and expansion steps in the refrigeration cycle. After a pre- 3,407,626 Patented Oct 29, 1968 determined volume of liquid has been vaporized within the chamber the outlet valve closes and a depressurizing means is actuated to reduce the pressure within the chamber and allow a fresh supply of liquid to enter the chamber. The liquid inlet valve closes after the liquid has reached a predetermined level in the chamber and the cycle of the pumping apparatus then recommences.

The novel features which are believed to be characteristic of the invention, both as to its organization and method of operation, together with further objects and advantages thereof will be better understood from the following description considered in connection with the accompanying drawing in which a presently preferred embodiment of the invention is illustrated by way of example. It is to be expressly understood, however, that the drawing is for the purpose of illustration and description only, and is not intended as a definition of the limits of the invention.

In the drawing:

FIGURE 1 is a partially schematic view in section of a thermal compressor utilized in the apparatus of the present invention at a first stage of operation;

FIGURE 2 is a view corresponding to FIGURE 1 of the thermal compressor at a subsequent stage of operation;

FIGURE 3 is a view corresponding to FIGURES 1 and 2 of the thermal compressor at a subsequent stage of operation;

FIGURE 4 is a view in section of a thermal compressor utilized in the present invention in the by-pass mode;

FIGURE 5 is a schematic view of an illustrative embodiment of a two-fluid refrigeration apparatus in accordance with the present invention;

FIGURE 6 is a schematic view of an illustrative embodiment of a single fluid refrigeration apparatus in accordance with the present invention;

FIGURE 7 is a pressure-enthalpy diagram of the operation of the apparatus of FIGURE 6; and

FIGURE 8 is a pressure-enthalpy diagram of the operation of the apparatus of FIGURE 5.

The refrigerating apparatus of the present invention utilizes a fluid pump as shown in FIGURES 1, 2 and 3 as more fully described in US. Letters Patent No. 3,285,001, filed concurrently herewith. The thermal pump of the present invention can be used to pump liquid in What is termed hereinafter the pump mode or to pump compressed gas or vapor in What is termed hereinafter the compressor mode. In the embodiments of FIG- URES 1 and 4 the thermally actuatedl pump is shown in the compressor mode. In connection with the embodiment of FIGURE 3, the compressor is shown in the self-actuated configuration as described more fully hereinafter while in FIGURE 4 the compressor is shown in the bypass mode as more fully described hereinafter and described and claimed in US. Patent No. 3,285,001.

Thus, referring to FIGURES 1 and 2 there is shown a thermally actuated fluid moving apparatus in the compressor mode. The compressor comprises in general a housing 30 defining a cylindrical chamber 31 within which is positioned a float valve as described more fully hereinafter. In the embodiment shown a fluid inlet port 34 communicates with the chamber 31 proximate the lower end wall 36 thereof. A gas outlet port 35 is positioned through the top wall of the housing in communication with the gas portion 31b of the chamber. Suitable inlet 37 and outlet 38 check valves are positioned in the inlet and outlet ports to allow the flow of fluid in the appripriate direction only. Suitable connectors 42 such as threaded nipples are provided for connecting input 33 and output 32 lines to the input and output ports. A heating element 45 is provided to supply heat. to the fluid within the chamber as described hereinafter. Any conveniently available source of heat can be utilized, however, for purposes of illustration an electrical resistance heater element is shown with a source of current 46.

From the upper portion of the chamber 31 a gas outlet port is provided through the top wall 49 of the housing and is operable from an open to a closed condition by the float valve 55. In the embodiment shown a gas outlet port 50, designated hereinafter as the depressurizing port 50, is located on the longitudinal axis of the chamber and has a tapered seat 51 divergent from the port diameter to the chamber. A suitable connector such as a threaded nipple 54 is affixed to the gas outlet port for connection of a depressurizing line 61.

The float valve 55 is so formed and of suitable material as to be buoyant in the liquid within the chamber 31. When the liquid is at a first predetermined level which is sufliciently high within the chamber, the fioat valve is raised to a position at which it closes the gas outlet port 50 and retains it closed until the liquid level drops to a predetermined second level. Thus, in the embodiment illustrated the float valve 33 has a cylindrical body portion 56 with an upwardy extending stem 57. The stem is vertically oriented in the orientation of the figures and is positioned on the center point of the upper surface of the float valve body 56 such that it is substantially coincident with the longitudinal centerline of the pump body. In its present embodiment the float valve is formed of nylon although other materials can be employed as will become apparent in connection with the description of the operation of the device and th function of the float valve. The length of the stem 57 is dependent upon the liquid level to be maintained in the chamber and length of stroke of the liquid-gas interface as described hereinafter. The stem has a valve element such as a metal ball affixed to its upper end and mateable with the valve seat 51. The outside diameter of the float valve body 56 is substantially less than the inside diameter of the chamber 31 and is determined by the volume of liquid to be displaced to obtain the necessary buoyancy forces upon the float valve. The float valve does not function in any way to prevent passage of liquid or gas from one portion of the chamber to the other. The function of the float valve 55 is solely to open and close the gas outlet port at a predetermined point in the cycle of the pump dependent upon the combination of forces acting upon the float valve. The forces operating upon the float valve and its function can most readily be seen in connection with the description of the operation of the invention in its most rudimentary form as shown in FIGURES 1 and 2.

Thus, referring now particularly to FIGURES 1 and 4 the compressor shown at a point in the cycle thereof where liquid is flowing into the pump compressor through the inlet check valve 37 which is open. The pressure within the compressor chamber 31 is at the pressure of the depressurizing line 61 since the depressurizing port 50 is open. The float valve 55 is at a position intermediate along its length of travel since as shown in FIG- URE 1 the liquid from the inlet has partially filled the chamber. The float valve is freely floating within the liquid to a depth of the liquid determined by the buoyancy forces on the float body 56. Thus, the float valve body will be submerged in the liquid to the depth indicated in the figures as a buoyancy line 73. As the liquid level continues to rise the float valve is carried upward until the valve element 61 on the valve stem engages and seats against the valve seat in the depressurizing line. Upon closing of the depressurizing line some immediate pressure differential exists between the chamber and the depressurizing line and a sudden closing and seating of the valve is obtained. A force is caused by this pressure differential across the float valve by reason of the fact that the upper transverse surface of the float valve against which the pressure within the chamber acts is less than the lower transverse surface by the amount of the area exposed to the depressurizing line as discussed more ful- 1y hereinafter. After the float valve is seated the liquid level ceases to rise within the chamber when the pressure of the fluid within the chamber slightly exceeds that of the fluid inlet line and causes the inlet check valve 37 to close. Since the heating element is energized throughout the operation of the pump, boiling of the fluid occurs during the latter part of the filling portion of the cycle. After the float valve closes the boiling of the liquid continues and produces vapor under pressure in the chamber above the liquid level. This pressure differential between the chamber pressure and the depressurizing line produces a differential force in the upward direction to retain the float valve in its seated position. For purposes of discussion the portion of the chamber filled with liquid is designated as 3112. The liquid surface is an interface which moves upwardly and downwardly within the chamber comparably to the piston face in a mechanical piston pump. As boiling occurs the pressure of the gas in portion 31b increases as does the pressure of the liquid in portion 31a.

Referring now to FIGURE 2, the liquid level is shown at or near its uppermost point immediately subsequent to opening of the outlet check valve 38. That is, after the inlet check valve and float valve have closed, boiling of the liquid continues and the pressure of the vapor and liquid in pressure balance continue to increase until the pressure of the liquid reaches and exceeds the pressure in the liquid outlet line. At this point the outlet check valve opens. When the outlet check valve 38 opens, a new pressure balance is established. Since the volume of gas generated by the boiling of liquid is substantially greater than the volume of liquid which is boiled to form such vapor the pressure of the vapor and the pressure of the liquid will remain constant at the outlet pressure from the compressor; i.e., the gas pressure in the outlet line. Consequently, gas will flow from the compressor through the outlet line at a substantially constant pressure and the pressure balance within the chamber 31 will remain substantially constant at this pressure as additional liquid is depleted and the liquid level moves downwardly. So long as the volumetric rate of gas generated equals the volume of gas passing from the compressor through the outlet line the pressure within the chamber will remain constant, gas will flow from the outlet at such pressure, and the liquid level will move downwardly. Referring now to FIGURES 2 and 3, as the liquid level within the pump chamber 31 decreases and passes below the upper surface 38 of the float valve body 56, the buoyancy force of the liquid acting upwardly upon the float valve begins to decrease. That is, the volume of liquid displaced by the float valve body 56 decreases as the liquid level passes downwardly beyond the buoyancy line .and a lesser volume of the float valve body is submerged in liquid. It can be seen, however, that the pressure forces operating upon the float valve remain constant, in that since the vapor pressure and liquid pressure are equal, the pressure exerted upwardly upon valve body 56 by the liquid acting upon the lower surface 65 of the valve body 56 is the same per unit area as the vapor pressure acting upon the upper surface 58 of the valve body. A pressure differential force thus exists since the upwardly directed forces due to pressure upon the float valve exceed the downwardly directed pressure forces.

It should be noted at this point that although the weight of the float valve is referred to herein only as a gravity force and the device is in a vertical orientation, other force inducing means such as acceleration, centrifugal force or magnetic force can be utilized to obtain the necessary force and pressure balances in accordance with the present invention when environmental conditions or other factors so require, as more fully discussed hereinafter.

Thus, referring now to FIGURE 4 there is shown an alternative embodiment of the thermal compressor of the present invention in which a bypass construction is utilized. The fluid chamber, float valve, liquid inlet line, and depressurizing line are identical in the illustrative embodiment to those of the embodiment shown in FIGURES 1 through 4 and are correspondingly identified. An outlet check valve 93 is positioned in the port 90 and the gas outlet line 91 is the high pressure side of the compressor at a pressure P A by-pass port 95 is in communication with the pump chamber 31 [at a position spaced downwardly :a substantial distance from the top wall of the chamber such that liquid or gas passes through the by-pass port dependent upon the liqui-d level within the chamber 31. The position of the bypass port is predetermined to determine the opening point of the depressurizing line as will become more apparent hereinafter. A restrictor means 97 such asan orifice of lesser diameter than the diameter of the bypass port, or a throttle valve, is positioned in the by-pass line to create a pressure differential under flow conditions be tween the chamber 31 and the by-pass line 93.

In FIGURE 4 the bypass embodiment of the thermal compressor in accordance with this invention is shown in the condition where the liquid level within the chamber 31 has risen to a point at which the buoyant float valve has been raised into sealing engagement between the valve stem element and the valve seat in the depressurizing port. At this point both the inlet check valve 37 and outlet check valve 93 are also closed. By means of the energized heating element the temperature of the liquid in the chamber 31 is raised to the point at which boiling commences and the pressure of the gas in the gas portion 31b of the chamber and also that of the fluid in the liquid portion 31a of the chamber are raised. The rise in pressure first seats the float valve firmly such that no leakage down the depressurizing line can occur. Secondly, a small quantity of liquid is by-passed from the compressor by flowing through the by-pass restrictor into the by-pass line and thence to the low pressure side of the compressor. Since the quantity of by-passed liquid is small due to the restriction in the line, the heating element continues to raise the pressure until the pressure within the compressor chamber 31 is equal to the pressure at the high pressure sideof the compressor, i.e., at the gas outlet line 91, at which time the outlet check valves 93 opens and allows gas to be pumped through the gas outlet line. The gas flows through the outlet line at substantially constant pressure and at a rate at which the volume of liquid flowing during a given time interval is equal to the volume of vaporized gas formed during that time interval. This process continues until the liquid level lowers by depletion of the liquid to the point at which the liquid level is at the by-pass port. At the level of liquid within the chamber 31 at which it passes beneath the bypass port a sudden pressure drop within the chamber occurs since the by-pass port is now in communication with the gas in the gas portion 31b of the chamber. Thus, gas rather than liquid is now forced by the high pressure through the by-pass orifice and by the proper design of the orifice size the pressure drop across the restrictor drastically decreases the pressure within the chamber 31. As this pressure suddenly decreases the float valve drops and the depressurizing line is opened. Thus, as described hereinbefore just prior to the pressure drop due to the escape of gas through the by-pass line the pressure forces operating upon the float valve to maintain it in the closed position are a combination of the buoyancy force of the liquid operating upon the float valve body together with the fluid pressure across the entire transverse lower surface of the valve body in the upper direction opposed by the weight of the float valve and the pressure of the chamber operating upon a lesser transverse surface area of the valve body. This pressure also maintains the outlet check valve open and the inlet check valve closed. When the pressure is decreased rapidly by the flow of gas through the by-pass line the pressure differential acting upon the upper and lower surfaces of the valve body terminates as soon as the pressure within the chamber reaches that within the depressurizing line. Thus, the weight of the float valve held in place by pressure forces causes it to fall, opening the depressurizing line and completing the depressurizing of the pump chamber to the low pressure point of the system. At this point the liquid head which feeds the inlet line to the pump chamber is sufficient to cause liquid to fill the pump chamber. As the liquid .fills the chamber the float valve is again buoyantlysupported and elevated to the position at which the depressurizing line is again closed and the pumping cycle is again initiated when the pressure within the chamber reaches the outlet line pressure. In this embodiment, when r the liquid level is below the buoyancy line of the float valve at the by-pass level the float valve will drop and the chamber 31 will be depressurized. The by-pass level can be made to occur at a point substantially beneath the float valve. That is, the material of which the float valve is formed and the configuration of the valve can be predetermined such that the ditferential pressure force upon the float valve is suflicient to overcome its weight and the float valve is maintained in the closed position although no buoyancy forces are exerted upon it. Thus, in the context of the previous description the drop line is not within the body of the float valve and the float valve will not fall to open the depressuring line: until the pressure drop in the chamber due to the passage of gas through the by-pass line is suflicient to reduce the upward differential force on the float valve below the Weight of the float valve.

Referring now to FIGURE 5, there is shown schematically a refrigerating apparatus in accordance with the present invention which utilizes conventional condensers, evaporators and expansion valves which are therefore not shown or described in detail. In the embodiment of FIG- URE 5 a two component fluid mixture is employed. In this embodiment one fluid component serves as a pumping fluid in the manner described hereinafter to produce the pumping action while the second fluid component provides the refrigerant action. The two fluids are designated hereinafter as the pump fluid and the refrigerant fluid. No molecular interaction between the two fluids can be allowed to exist. Freon 114 as the pumping fluid and Freon 12 as the refrigerant fluid are exemplary of the fluids which can be intermixed. Further illustrative combinations of fluids are given hereinafter. In accordance with the present invention a thermal compressor as shown in FIGURES 1, 2 and 3 is utilized as the prime mover for a refrigeration cycle which utilizes the standard compression-condensation expansion steps in the cycle. Accordingly, in FIGURE 5 there is shown schematically a refrigerant condenser 100 and evaporator 101 with an expansion valve 102 positioned therebetween in the fluid line 103 all in a manner well known to the art. The condenser 100 is positioned exteriorly of the space to be cooled, i.e., at the high temperature side of the system while the evaporator is within the space to be cooled, i.e., at the low temperature side of the system. A second condenser 104 serves as a condenser for the pumping fluid and is positioned at the low pressure side of the thermal compressor with inlet 105 to the condenser 104 connected to the depressurizing line 61 of the thermal compressor and the outlet 106 from the condenser 104 connected to the liquid inlet line 33 to the thermal compressor. Thus, the thermal compressor cycles as previously described. During the liquid inlet portion of the cycle a mixture of the refrigeration and the pumping fluids are admitted to the chamber 31 of the thermal compressor. After the float valve has seated and closed the depressurizing line a supply of gas is conducted from the thermal compressor through the gas outlet line 91 to the refrigerant condenser 100. The high pressure gas expelled from the thermal compressor is admixed refrigeration fluid and pumping fluid in the gaseous state. The two fluids are so chosen that in passing through the condenser 100 both the refrigeration fluid and pumping fluid are liquified while in passing through the expansion valve only the refrigeration fluid is again partially transformed to the gaseous state. Accordingly, through the evaporator 101 only the refrigeration fluid takes on appreciable heat by vaporization. The intermixed pumping fluid in liquid state and refrigeration fluid in gaseous state are then conducted to the fluidinlet of the compressor for the commencement of another cycle. The refrigerant fluid and pumping fluid released from the chamber 31 when the depressurizing line 61 is opened are conducted through the pump condenser 104 where the pumping fluid only is liquified after which the admixed pumping liquid and refrigerant gas are also conducted to the compressor inlet line 33. In practice very little refrigerant fluid passes through the pump cycle as shown hereinafter. The steps in the above cycle can be clarified by reference to FIGURE 8 in which the thermodynamic cycle is shown in separate pressureenthalpy diagrams for the pumping fluid and refrigerant fluid. The pumping fluid cycle is shown as 8a while the refrigerant fluid cycle is shown as 811. The intermixed fluids in gaseous form are pumped from the thermal compressor 30 through a check valve 110 at high pressure and high temperature. In the illustrative embodiment utilizing Freon 12 and Freon 114 the outlet temperature of the intermixed gases from the thermal compressor is 180 F. and the pressure is 140 p.s.i. which is the evaporation temperature and pressure of the pumping fluid Freon 114. Since the evaporating temperature of Freon 12 is 40 F. and the evaporating pressure is 50 p.s.i. the refrigerant fluid component of the intermixed gases leaving the thermal compressor is superheated. Thus, referring to FIG- URES 5, 8a and 8b, the compressor outlet line 91 is at the high pressure-high temperature point of the cycle at the pressure P as designated hereinbefore. In FIGURES 8a and 8b the point of the cycles of the two fluids at the outlet line 91 are shown as 91p on the pumping fluid diagram and 91r on the refrigerant fluid diagram. Thus it can be seen that at the outlet pressure of 140 p.s.i. and temperature of 180 F. required for the vaporization of the F114 pumping fluid, the pumping fluid is at the vapor stage 91p while the refrigerant fluid is a superheated vapor at 912 corresponding to the high temperature T of 180 at the high pressure P of 140 p.s.i. It can be seen that the pressure in the system of FIGURE 5 is at P through the condenser designated the refrigerant condenser 100. The temperature of the condenser is near ambient and suflicient to reduce the intermixed gases to their condensing temperature which is 106 F. for the refrigerating fluid and 180 for the pumping fluid. Thus, as the intermixed fluids pass through the condenser 100 from line 91 to the inlet 103 to the expansion valve 102 they pass along thecondensation line in FIGURES 8a and 81) from 91 to 103 and are reduced in temperature from 180 F. to approximately 106 F. designated ambient temperature T It can be seen that at point 103 in FIG- URE 8b the refrigerant fluid has reached the condensation point of the pressure-enthalpy chart while the pumping fluid has been already liquified and subcooled from T to T i.e., from 180 F. to 106 F. Thus, upon entering the expansion valve the intermixed fluids are both liquid and are at a pressure of 140 p.s.i. and a temperature of 106 which is the condensation pressure and temperature of the Freon 12.

Across the expansion valve the pressure of the intermixed fluids is dropped to the low pressure of the systern P As shown in FIGURE 8b in dropping from P to P the refrigerant fluid does so at constant enthalpy from point 103 to point 101a at the inlet to the evaporator 101. At the inlet to the evaporator the fluids are therefore at a pressure of 50 p.s.i. and a temperature of approximately 40 F. which is the evaporation pressure and temperature of the Freon 12. As shown at 101a in FIGURE 81), the Freon 12 is partially liquid and partially gas at this low temperature and in passing through the evaporator 101 evaporates at the constant temperature of 40 F. and absorbs heat from the surrounding spaces. In passing through the evaporator the Freon 12 which is part liquid-part gas at 101a as shown in FIGURE 8b vaporizes along the constant pressure-constant temperature line from 101a to 10117 at which point it .is substantially all gas.

During the same period of the cycle the pumping 'fluid with which the refrigerant fluid is intermixed has been subcooled as a liquid from the temperature at which it left the condenser to the refrigerant temperature, i.e., from point 103 in FIGURE 8a to 101a at which it is at the temperature of 40 F. and pressure of 50 p.s.i. as described above. In'passing through the evaporator from the point 101a to 10111 of thethermodynamic cycle the pumping fluid does not change in pressure or temperature since it is intermixed with the refrigerant fluid which evaporates at constant pressure and temperature of 40 F. while the pumping fluid at this temperature is substantiallybelow its evaporating temperature. Thus, points 101a and 10111 are coincident in FIGURE 8a. Accordingly when the intermixed refrigerant and pump ing fluids are in the inlet line 33 to the thermal compressor the refrigerant fluid is an entrained gas in the pumping liquid and both are at a temperature and pressure of 50 p.s.i. and 40 F. respectively. It can be seen that the point of the thermodynamic cycle at which the fluids leave the evaporator is dependent in part upon the temperature of the cooled spaces surrounding the evaporator and the heat transfer extent of the evaporator. Thus, the point 101b may occur in the evaporator or at the thermal compressor.

In the thermal compressor the temperature and pressure of the intermixed fluids are raised, as discussed hereinbefore, after suflicient liquid has been admitted to the thermal compressor to seat the float valve. The temperature of the liquid within the thermal compressor is raised by a heat source until it is vaporized at a constant pressure of p.s.i. Thus, within the thermal compressor the refrigerant fluid is raised from the point indicated as 101b in FIGURE 8b to the point indicated as 91 which is the outlet temperature and pressure of the refrigerant fluid from the thermal compressor. The refrigerant fluid is thus raised from a temperature of 40 F. to F. and from a pressure of 50 p.s.i. to 140 p.s.i. As described hereinbefore the pumping fluid upon entering the thermal compressor is at the point indicated as 10112 in FIGURE 8a which corresponds to point 101a, the pumping fluid at this point being a subcooled liquid. Accordingly, in the thermal compressor the pumping fluid in the liquid state is raised to the point at which it begins to evaporate at the lower pressure. That is, as shown in FIGURE 8a, as heat is supplied to the subcooled liquid its temperature rises along the line from 101b to 10112 at which the pressure has increased slightly and the temperature has been raised to the point at which some evaporation at the lower pressure commences. The temperature risecontinues until, the higher pressure T of 140 p.s.i. is reached at which time a constant pressure evaporation process is carried on in the thermal compressor from the point indicated as 91 to 91 in FIGURE 8a. Upon leaving the thermal compressor for thecommencement of another cycle as described above the pumping fluid and refrigerant. fluid are .both in the gaseous state at 180 F. and 140 p.s.i.

Since the refrigerant fluid has a lower boiling pressure and temperature than does the pumpingfluid the first gas pumped from the compressor will be the refrigerant fluid and bythetirne the liquid level of the liquid Within the thermal compressor has dropped to the point at which the float valvefalls to openthedepressurizing line. All I of theprefrigerant fluid has. been expelled in. compressed condition from the. thermal: compressor. Accordingly,

through the depressurizing outlet the gas which is-released from, the thermal compressorto-depress-urize the compressor and reduce the internal pressure thereof to the value of P as described hereinbefore is substantially all pumping fluid. Thus, upon opening of the depressurizing port pumping fluid in the gaseous form escapes from the thermal compressor and is conducted through the pumping fluid condenser 104 where it is reduced in temperature to the ambient. Referring to FIGURE 8a the pumping fluid in the pumping fluid condenser portion of the apparatus is throttled from the high pressure at point 91 of the pressure enthalpy diagram to the low pressure at point 105 after which it is condensed from point 105 to 106 through the pumping condenser. It then joins the intermixed pumping and refrigerant fluid in the inlet line 33 to the thermal compressor and enters the thermal compressor to commence another cycle.

Examples of other combinations of refrigerant fluids and pumping fluid such as Freon ll-Freon 21 with a P =80 p.s.i.a., P =49 p.s.i.a., T =112 F., T =144 F., T =180 F., will be apparent to those skilled in the art in view of the foregoing.

Referring now to FIGURE 6 there is shown a schematic diagram of an alternative embodiment of the present invention in which a single fluid is utilized rather than intermixed pumping and refrigerant fluids as previously described. In this embodiment a condenser 100, expansion valve 102 and evaporator 101 of the type well known to the art and previously described are utilized. A thermal compressor 30 is again used in accordance with the present invention and is shown in the bypass con figuration. A refrigeration compressor 120 is shown schematically since it can be one of many types of compressors in which a fluid under pressure is utilized as the energy source for moving the compressor means which in turn compresses a gas. For example, a differential free piston compressor can be employed as can a simple turbo-compressor in which gas under pressure moves the turbine 120 after which the gas which has provided the moving force for the turbine is exhausted at a lower pressure and temperature and the gas which is being compressed is increased in pressure and temperature. Generally, means may be regarded as a converter for receiving a high pressure fluid and a low pressure fluid and for supplying an intermediate pressure fluid. Thus as shown schematically in FIGURE 6 a high pressure fluid line 91 is conducted to the turbine 120 at the periphery thereof which comprises blades, buckets or the like. The fluid under pressure drives the turbine 120 and the compressor 121 attached thereto and is 6X- hausted at reduced pressure P through outlet line 92. Conversely, low pressure fluid P is admitted to the compressor 121 at a radially inward inlet 123 and compressed and exhausted at increased pressure and temperature at compressor outlet line 92a. In the embodiment of FIGURE 6 Freon 12 can be utilized as an illustrative fluid for the apparatus of the present invention. The pressure enthalpy diagram of the illustrative embodiment is shown in FIGURE 7 for purposes of discussion. Referring now to FIGURES 6 and 7 Freon 12 is compressed and expelled from the thermal compressor in the manner previously described. In FIGURE 7 it can be seen that the fluid is expelled in gaseous form from the thermal compressor at a pressure of approximately 280 p.s.i. and a temperature of 160 F. The high pressure high temperature gas is transmitted to the refrigeration compressor along outlet line 91. The condition of the gas is thus shown at point 91 on the pressure enthalpy diagram of FIGURE 7. The gas under a pressure of 280 p.s.i. and at a temperature of 160 F. thus enters the turbo-compressor at the driving portion thereof, i.e., the turbine at inlet 123 and drives the turbine within the compressor until a point of movement is reached at which the gas is exhausted from the turbine. This is shown schematically as exhaust line 92 from the turbo-compressor. In performing the work of compression the high pressure gas is reduced in pressure and temperature from point 91 to the point designated as 92. in FIGURE 7 which in the illustrative embodiment is a temperature of approximately F. and a pressure of 130 p.s.i. The pressure of 280 p.s.i. is the high pressure of the system while the pressure of 130 p.s.i. is the intermediate pressure of the system. The low pressure point of the system is subsequent to the expansion and cooling of the gases but with reference to the refrigeration compressor occurs at the inlet point 123 of the gas to be compressed at the inlet line 115. Thus, as will be seen hereinafter the gas entering the compressor which is to be raised in temperature and pressure enters at a pressure of 35 p.s.i. and a temperature of 20 at point 123 in FIGURE 7. After being compressed it is exhausted from the compressor at the intermediate pressure of 130 p.s.i. and a temperature of 100". Thus, the low pressure gas is compressed to the intermediate pressure at point 92 of FIGURE 7 and mixed with the gas which has been raised in pressure such that in line 116 the pressure and temperature of the intermixed gases are 130 p.s.i. at 100 F. The compression of the gas in the compressor therefor occurs between the points indicated as 123 and 92 in FIGURE 7. As shown in FIGURE 6 the depressurizing port of the thermal compressor is also connected through the depressurizing outlet line to a junction point with the compressor line 116 which point is identified as 117 in FIGURE 6. Thus, at point 1.17 all of the gas in the system is at a pressure of 130* p.s.i. and a temperature of 100 F. The point 117 corresponds to the inlet to the condenser 100 and the gas is conducted through such condenser and reduced in pressure and temperature along the line designated! as 92 to in FIGURE 7. Along this line the gas is condensed and at the outlet 105 from the condenser is: at a pressure of 130 p.s.i. and a temperature of 100 and is liquid in form. Prior to the inlet to expansion valve 102 a junction point in the intermediate pressure line is provided by means of which a substantial portion of the liquid at 130 p.s.i. is conducted back to the inlet 33 to thermal compressor 30. Thus, with respect to this fluid the cycle is completed when the liquid re-enters the thermal compressor. This returned portion of the liquid when returned to the thermal compressor is heated within the thermal compressor 30 along the line shown as 105 to 30 in FIGURE 7 and expelled from "the compressor at constant pressure along line 3091 of FIGURE 7. The remainder of the fluid passes through the expansion valve where it is reduced in pressure and thus becomes a mixture of liquid and gas at the low pressure of 35 p.s.i. and the temperature of 20 F. Thus, this expansion of the fluid occurs along the line 105 to 101 in FIGURE 7. The gas then passes through the evaporator 101 where condensation occurs to extract the heat from the surrounding spaces. Thus, along line 101-123 in FIGURE 7 the evaporation of the liquid passing through the evaporator occurs until at point 123 the liquid has substantially all been evaporated and re-enters the refrigeration compressor 121 at the low pressure of 35 p.s.i. and low temperature of 20 F. The cycle then recommences. The by-pass line of the thermal compressor in the mode shown in FIGURE 6 also allows gas at the intermediate pressure to escape from the thermal compressor and such by-pass gas is conducted to the mixing point 117 to proceed through the cycle with the balance of the fluids. It should be noted that the path from the thermal compressor 30 to converter means 120 to condenser 100 and back to thermal compressor 30 (represented by loop 30, 91, 92 and 105 on FIGURE 7) is the power cycle of the system. The path from converter means 120 to condenser 100 to evaporator 101 and back to converter means 120 (represented by a loop 92, 105, 101 and 123 of FIGURE 7) is the refrigeration cycle of the system.

It will be apparent from the foregoing that other embodiments of the present invention can be employed in accordance with present invention. For example, liquid under pressure from the thermal compressor operated in the pumpi-ngmode wherein liquid rather than gas is pumped therefrom can be utilized as the means for driving the refrigerant compressor.

Thus, the present invention provides an improved refrigeration method and apparatus.

What is claimed is:

1. A refrigeration system comprising:

a housing defining a fluid chamber;

a fluid outlet from said chamber;

a fluid inlet to said chamber;

a depressurizing outlet from said chamber;

means for vaporizing liquid within said chamber;

liquid volume dependent means for closing said depressurizi-ng outlet at a first liquid volume and for retaining said depressurizing outlet closed until sufficient of said liquid has been vaporized to reduce the volume of said liquid to a substantially lesser liquid volume;

means in combination with said fluid inlet for opening said inlet at said second liquid volume and closing said inlet at said first liquid volume;

means in combination with said fluid outlet for closing said outlet at said second liquid volume and opening said outlet at said first volume;

said fluid being a refrigerant fluid;

refrigerant compressor means adapted to be driven by said fluid under pressure;

a condenser;

means for conducting said fluid under pressure to said compressor, means for conducting the fluid exhausted from said compressor to said condenser inlet;

an evaporator;

an expansion value interposed between said condenser and evaporator; means for coupling said depressurizing outlet to a depressurizing pressure whe opened;

means for conducting a portion of said fluid from said condenser to said fluid inlet; and,

means for conducting a portion of said fluid through said expansion valve, evaporator and to said compressor.

2. A device for moving fluid under pressure comprising:

a housing defining a fluid chamber;

a fluid inlet to said chamber;

a fluid outlet from said chamber;

a depressurizing vapor outlet port from said chamber;

heating means for vaporizing liquid of said chamber whereby said chamber contains a volume of liquid and a volume of vapor, said heating means being adapted to supply heat sufficient to vaporize said liquid and raise the pressure of said fluids within said chamber to a high pressure P from a lower pressure P a float valve adapted to be buoyantly supported by said liquid and moved to a first position by increase of the liquid volume in said chamber to a first predetermined liquid level;

said float valve being so constructed and arranged as to provide a valve element portion adapted to seat in and close said depressurizing port in said first position of said float valve, the pressure at the exterior of said port to which said valve element portion is exposed being at a pressure P less than P such that in said first position said float valve is urged toward said first position by the buoyancy force of said liquid and the force exerted by the pressure differential between P and P acting upon said float valve, said float valve being urged away from said first position by the weight thereof whereby said float valve is moved to a second predetermined position by decrease of the liquid volume to a second level at which the buoyancy force and differential force are less than the weight force of said float valve, said depressurizing port being opened by movement of said float valve from said first to said second position; means for coupling said depressurizing outlet to a depressurizing pressure when opened;

outlet valve means adapted to open said fluid outlet when said fluid pressure within said chamber reaches the higher pressure--P -"-andclose said fluid outle when said pressure is less than P inlet valve means adapted to open 'said'liquid inlet when said chamber pressure is less than P and close said liquid inlet when said chamber pressure exceeds 2;

said fluid being a refrigerant fluid;

refrigerant compressor means adapted to be driven by said fluid under pressure;

a condenser;

means for conducting said fluid under pressure to said compressor, means for conducting the fluidexhausted from said compressor to said condenser inlet;

fluid inlet means to a driven portion of said compressor, fluid outlet means from said driven portion of said compressor, means for conducting said fluid from said driven portion outlet means to said condenser inlet;

an evaporator;

an expansion valve interposed between said condenser and evaporator; and

means for conducting a portion of said fluid from said condenser to said liquid inlet to said thermal compressor;

means for conducting the balance of said fluid through said expansion valve, evaporator and to the inlet of said driver portion of said refrigerant compressor.

3. A device for moving fluid under pressure comprising:

a housing defining a closed fluid chamber;

a liquid inlet to said chamber;

a vapor outlet from said chamber;

a depressurizing vapor outlet port from said chamber; means for coupling said depressurizing' outlet to a depressurizing pressure whenopened;

heating means for vaporizing liquid within said chamber whereby said chamber contains a volume of liquid and a volume of vapor, Said'heating means being adapted to supply heat suflicient to vaporize said liquid andiraise the pressure of said fluids within said chamber to a high pressure P from a lower" said float valve being so constructed and arrangedas to provide a valve element portion adapted to seat in and close said depressurizing port in said first position of said float valve, the pressure at the exterior of said port to which said valve element portion is exposed being at a pressure P less than P such that in said first position said float valveis urged toward said first position by the buoyance force of said liquid and the force exerted by the pressure differential between P and P acting upon said float valve, said float valve being urged away from said first position by the weight thereof whereby said float valve is moved to a second predeter-,,

float valve, said depressuriz-ingport being openedv by movement of said float valve from said first to said second position;

outlet check .valve means adapted to open, said vapor outlet when said fluid pressure within saidchamber reachesthe'higher pressure P 'and close said flui outletwhensaidpressureislessthanP inlet check valve means adapted to open said liquid 13 inlet when said chamber pressure is less than P and close said liquid inlet when said chamber pressure exceeds P said fluid being a refrigerant fluid;

refrigerant compressor means adapted to be driven by said fluid under pressure;

a condenser;

means for conducting said fluid under pressure to said compressor, means for conducting the fluid exhausted from said compressor to said condenser inlet;

fluid inlet means to a driven portion of said compressor, fluid outlet means from said driven portion of said compressor, means for conducting said fluid from said driven portion outlet means to said condenser inlet;

an evaporator;

an expansion valve interposed between said condenser and evaporator; and,

means for conducting a portion of said fluid from said condenser to said liquid inlet to said thermal compressor;

means for conducting the balance of said fluid through said expansion valve, evaporator and to the inlet of said driver portion of said refrigerant compressor.

4. A refrigeration system comprising:

a chamber having an inlet and an outlet;

a refrigerant within said system, said refrigerant existing in liquid and gaseous states in said system, said proportion of liquid and gas in said chamber varyin a heating means for heating said refrigerant, said means associated with said chamber to vary the proportion of gas and liquid therein;

means for opening said inlet and closing said outlet when said refrigerant liquid is below a first level and for opening said outlet and closing said inlet when said refrigerant liquid is at a second and higher level;

said heating means increasing the proportion of gas in at least said chamber and increasing said pressure associated therewith when said outlet is closed to create a higher pressure refrigerant;

a condenser;

an evaporator;

an expansion valve disposed between said condenser and evaporator and coupled thereto;

a converter means for receiving refrigerant from said evaporator and coupled thereto, for receiving high pressure refrigerant, and for providing refrigerant with a pressure between the pressure of said refrigerant from said evaporator and said higher pressure refrigerant;

means for coupling a portion of said intermediate pres sure refrigerant from said converter to said condenser; and

means for effectively coupling said condenser to said chamber.

5. The structure recited in claim 4, wherein said means for heating heats refrigerant passing from said chamber and vaporizes such refrigerant to lower the level of liquid refrigerant in said chamber;

and said means for opening said inlet and closing said outlet is a liquid level dependent means.

6. The structure recited in claim 5, wherein said liquid level dependent means is a float.

7. The structure recited in claim 5, wherein said high pressure refrigerant is provided by said heating means and said chamber for at least a portion of each cycle of said refrigeration system operation.

8. A refrigeration system comprising:

a housing defining a fluid chamber;

a fluid outlet from said chamber;

a fluid inlet to said chamber;

a depressurizing outlet from said chamber;

a means for coupling said depressurizing outlet to a depressurizing pressure when opened;

heating means for vaporizing liquid of said chamber;

liquid leveldependent means for closing said depressurizing outlet at a first liquid level and for retaining said depressurizing outlet closed until sufficient of said liquid has been vaporized to reduce the level of said liquid to a second lower level;

means in combination with said fluid] inlet for opening said inlet in the proximity of said second liquid level and closing said inlet in the proximity of said first liquid level; means in combination with said fluid outlet for closing said outlet in the proximity of said second liquid level and opening said outlet in the proximity of said first liquid level;

said fluid being a refrigerant;

converter means for receiving fluid from said chamber and heating means at a first pressure and from said evaporator at a second and lower pressure and for providing refrigerant at an intermediate pressure;

a condenser;

means for conducting said high pressure fluid to said converter means;

means for conducting the fluid from said converter means to said condenser;

an evaporator;

an expansion valve interposed between said condenser and said evaporator and coupled thereto;

means for conducting a portion of said fluid from said condenser to said fluid inlet;

and means for conducting a portion of said fluid from said evaporator to said converter means.

References Cited UNITED STATES PATENTS 653,171 7/1900 Coleman 62-88 2,991,632 7/1961 Rogers 62-467 3,153,442 10/1964 Siluern 62-467 MEYER PERLIN, Primary Examiner. 

1. A REFRIGERATION SYSTEM COMPRISING: A HOUSING DEFINING A FLUID CHAMBER; A FLUID OUTLET FROM SAID CHAMBER; A FLUID INLET TO SAID CHAMBER; A DEPRESSURIZING OUTLET FROM SAID CHAMBER; MEANS FOR VAPORIZING LIQUID WITHIN SAID CHAMBER; LIQUID VOLUME DEPENDENT MEANS FOR CLOSING SAID DEPRESSURIZING OUTLET AT FIRST LIQUID VOLUME AND FOR RETAINING SAID DEPRESSURIZING OUTLET CLOSED UNTIL SUFFICIENT OF SAID LIQUID HAS BEEN VAPORIZED TO REDUCE THE VOLUME OF SAID LIQUID TO A SUBSTANTIALLY LESSER LIQUID VOLUME; MEANS IN COMBINATION WITH SAID FLUID INLET FOR OPENING SAID INLET AT SAID SECOND LIQUID VOLUME AND CLOSING SAID INLET AT SAID FIRST LIQUID VOLUME; MEANS INCOMBINATION WITH SAID FLUID OUTLET FOR CLOSING SAID OUTLET AT SAID SECOND LIQUID VOLUME AND OPENING SAID OUTLET AT SAID FIRST VOLUME; SAID FLUID BEING A REFRIGERANT FLUID; REFRIGERANT COMPRESSOR MEANS ADAPTED TO BE DRIVEN BY SAID FLUID UNDER PRESSURE; A CONDENSER; MEANS FOR CONDUCTING SAID FLUID UNDER PRESSURE TO SAID COMPRESSOR, MEANS FOR CONDUCTING THE FLUID EXHAUSTED FROM SAID COMPRESSOR TO SAID CONDENSER INLET; AN EVAPORATOR; AN EXPANSION VALUE INTERPOSED BETWEEN SAID CONDENSER AND EVAPORATOR; MEANS FOR COUPLING SAID DEPRESSURIZING OUTLET TO A DEPRESSURIZING PRESSURE WHEN OPENED; MEANS FOR CONDUCTING A PORTION OF SAID FLUID FROM SAID CONDENSER TO SAID FLUID INLET; AND MEANS FOR CONDUCTING A PORTION OF SAID FLUID THROUGH SAID EXPANSION VALVE, EVAPORATOR AND TO SAID COMPRESSOR. 